Advantages
High available energy at turbine
Good performance at low speed and load
Good turbocharger acceleration

Disadvantages
Poor turbine efficiency with one or two cylinders per turbine entry
Poor turbine efficiency at very high ratings
Complex exhaust manifold with large numbers of cylinders
Possible pressure wave reflection problems (on some engines)

These turbochargers are characterized by having axial flow, single stage, turbines and are fitted to the majority of large industrial and marine engines, both four- and two-stroke. The duty cycles of these engines are more arduous than that of automotive engines and they tend to spend much more of their operating time at high load. Furthermore the consequences of failure are more serious, particularly on a marine engine. As a result, although every attempt is made to keep the designs simple, the primary objectives are a very high level of reliability, high efficiency and versatility to cover a great range of engine types and sizes at reasonable cost. However, design variations from one manufacturer to another are greater than is the case with smaller turbochargers.

Figure 2.10 is a cross-section of a typical large turbocharger, with a radial flow compressor and axial flow turbine. The compressor impeller is made in two separate parts, the inducer and main part of the impeller. The inducer is usually machined from a steel casting or an aluminium forging, and is splined or keyed to the shaft. The impeller is machined from an aluminium forging except for very high pressure ratio requirements when titanium is used due to its superior high temperature properties. The advantage of the two piece compressor is ease of machining, but an additional benefit is some impeller vane damping provided by friction at the inducer-impeller contact surfaces. Compressor diffusers are vaned for high efficiency.

The turbine disc is machined either as an integral part of the shaft or is shrunk on to the shaft. The rotor blades may be cast, forged or machined from a high temperature creep-resistant steel such as Nimonic 8OA or 90. Welded joints or ‘fir-tree’ roots are used to fix them to the disc, the latter design being more common on high pressure units since they provide a degree of vibration damping and allow a wider selection of blade and disc materials to be considered. Additional vibration damping can be provided by wire lacing the blades. The turbocharger manufacturer will offer a range of ‘trims’ or flow capacities with each basic design of turbocharger by varying blade (stator and rotor) height and stator blade angle.

A disadvantage of the axial flow turbine is that it complicates the design of the gas inlet and outlet. The gas inflow section is particularly important hence this is usually located on the end, allowing generous curvature in the inlet ducts to the stator blades for minimum flow distortion and loss. The turbine exit duct acts merely as a collector, hence a compact design can be used, minimizing turbocharger length. However, a recent trend is to utilize some exhaust diffusion to increase turbine expansion ratio and power output.

Most of the larger turbochargers in this class have outboard rolling element bearings (i.e. outside the compressor and turbine, Figure 2.10), with their own oil supply, and resilient mountings to prevent brinelling. The advantages of this are stable shaft mounting and low dynamic loads due to the wide bearing spacing, small bearing diameter, low rolling resistance and good access for bearing maintenance. The use of separate oil supplies for the turbocharger and engine enables a lower viscosity oil to be used, further reducing bearing friction. Low pressure ratio turbochargers use simple rotating steel discs, partially immersed in the oil, to pick up and deliver the oil to the bearings, but with higher bearing loads and speeds, gear pumps are used to spray oil on the bearings. Plain or sleeve bearings are sometimes available as an option and are preferred for durability although their frictional losses are greater.

Turbocharger design is simpler with inboard bearings since this gives greater freedom to design low loss intake ducts. Fewer components are required and the turbocharger is shorter, lighter and cheaper as a result (Figure 2.11). The disadvantage is a less stable bearing system and higher bearing loads. Fully floating sleeve and multi-lobe plain bearings are used, with well damped mountings for stability; the rotors must still be carefully balanced. Relative to rolling element bearings, higher oil pressure and greater oil flow rates are required and the combination of large diameter and width means that frictional losses are greater.

With either bearing system, the turbine outlet casing is the main structure to which the other components are bolted, and incorporates mountings to the engine. The casing is usually water cooled. Bolted to it is the water cooled turbine inlet casing, incorporating the bearing housing (for outboard bearings) and its oil reservoir. Single, two, three and four entry turbine inlets are available, manufactured from high grade cast iron. Between turbine inlet and outlet casings, provision is made for mounting the turbine stator nozzle ring. The compressor inlet and outlet casings are aluminium alloy castings.

The compressor inlet casing incorporates webs to support the bearing housing if outboard bearings are used. These webs must be carefully designed to be far enough away from the impeller to avoid impeller vane excitation. The casing also houses a combined air filter and silencer on most larger turbochargers (Figure 2.10). Sound waves originating at the compressor intake are reflected and reduced in intensity by baffles lined with sound absorbing material.

Turbochargers designed for small industrial and marine engines, though larger than those of large truck engines, are similar in concept to the automotive turbochargers described above. Radial flow compressor and turbines are used, with an inboard bearing arrangement. Apart from the larger size, they are required to have greater durability and higher efficiency. Thus the designs are usually more complex and expensive.

Engines designed for these applications operate over a smaller speed range than truck engines, and at greater b.m.e.p., hence higher compressor pressure ratio. It follows that the flow range required from the compressor is smaller, hence vaned diffusers are used. Vaned turbine stator nozzles are also used. This results in higher design point compressor and turbine efficiency. A range of diffuser nozzle angles and turbine stator blade angles are available for matching a basic turbocharger to a particular engine.

The maximum size is governed by precision casting limitations for the radial flow turbine rotor, currently about 300 mm, although most units in this class are smaller. Turbine housings are simple volutes designed to deliver the flow evenly around the circumference of the stator nozzle ring, the latter generating the design gas flow angle at rotor inlet. The turbine housings are supplied in uncooled or water cooled form. Although cooling is undesirable thermodynamically, it is sometimes required for safety reasons due to the potential danger of hot exposed surfaces in small engine rooms.

Bearings are of similar design to those of automotive units, except that clearances, relative to turbocharger size, are smaller. Sometimes cooling air is bled from the compressor to the rear of the turbine hub and bearing area. This also helps prevent exhaust gas leaking down the back of the turbine wheel and reaching the bearings. These techniques help keep the hot end bearing cool, preventing serious oil oxidation deposits. Like the smaller units previously described, the lubricating oil system of the engine is also used for the turbocharger. Since bearing clearances are smaller, rotor movements are small and conventional labyrinth oil seals can be used at the compressor and turbine ends of the rotor shaft.

Turbochargers of this type are made in relatively small numbers, by batch production, hence their cost is high relative to automotive units.

The compressor impeller is an aluminium alloy (LM- 16-WP or C-355T61) investment casting, with a gravity die-cast aluminium housing (LM-27-M). The design of the impeller is a compromise between aerodynamic requirements, mechanical strength and foundry capabilities. To achieve high efficiency, and minimum flow blockage, very thin and sharp impeller vanes are required, thickening at the root (impeller hub) for stress reasons. It is common practice to use splitter blades that start part way through the inducer, in order to maintain good flow guidance near the impeller tip without excessive flow blockage at the eye. Until recently the impeller vanes have been purely radial so that blades were not subjected to bending stress. However most recent designs incorporate backswept blades at the impeller tip since this has been shown to give better flow control and reduces flow distortion transmitted through from impeller to diffuser.

Automotive turbocharger compressor impeller, with splitter blades.

Typical design point pressure ratios fall in the range of 2 to 2.5:1, requiring impeller tip speeds of 300 to 350 m/s, hence small units of typically 0.08 m tip diameter rotate at 72 000 to 83 000 rev/min. In order to match wide differences in air flow requirements from one engine to another, a range of compressor impellers is available to fit the same turbocharger. These will be produced from one or two impeller castings, but with different tip widths and eye diameters generated by machining as shown in Figure 2.6, and matched with appropriate compressor housings. Usually up to ten or more alternative ‘trims’ are available but since the impeller tip diameter is unchanged and the hub diameter at the impeller eye is fixed by the shaft diameter, the flow passage variations alter the efficiency as well as flow characteristics of the impeller.

The compressor can be a loose or slight interference fit on the shaft, clamped by the compressor end nut. Impellers of most turbochargers are balanced before assembly onto the shaft, so that components can be interchanged without rebalancing.

Vaneless diffusers are used on all except very high pressure ratio compressors. Relative to the alternative vaned designs, the vaneless diffuser is slightly less efficient due to a longer gas flow path and poorer flow guidance, but has a substantially wider range of high operating efficiency. This is important in truck and passenger car applications where engine speed, and therefore mass flow range, is large. The volute acts not merely as a collector of air leaving the diffuser, but is usually designed to achieve a small amount of additional diffusion in its delivery duct. Generally the volute slightly overhangs the diffuser in order to reduce the overall diameter of the turbocharger. The volute and impeller casing are invariably formed as a single component.

Turbochargers in this class are used for passenger car diesel engines rated at 45 kW upwards to larger special heavy truck and construction vehicles rated at up to 600 kW. The most important design factors are cost, reliability and performance. To keep cost low, the design must be simple, hence a single stage radial flow compressor, and a radial flow turbine are mounted on a common shaft with an inboard bearing system. This arrangement simplifies the design of inlet and exhaust casing and reduces the total weight of the turbocharger.

The vast majority of current diesel engines operate on the four stroke principle in which combustion occurs only every other revolution, again in the region of top dead centre (TDC), and with the intermediate revolution and its associated piston strokes given over to the gas exchange process. In practice the exhaust valve(s) open well before bottom dead centre (BDC) following the expansion stroke and only close well after the following top dead centre (TDC) position is reached. The inlet valve(s) open before this latter TDC, giving a period of overlap between inlet valve opening (IVO) and exhaust valve closing (EVC) during which the comparatively small clearance volume is scavenged of most of the remaining products of combustion. Following completion of the inlet stroke, the inlet valve(s) close well after the following bottom dead centre (BDC), after which the ‘closed’ portion of the cycle, i.e. the sequence compression, combustion, expansion, leads to the next cycle, commencing again with exhaust valve opening (EVO).

The main advantages of the four-stroke cycle over its two stroke
counterpart are:

(a) the longer period available for the gas exhange process and the separation of the exhaust and inlet periods— apart from the comparatively short overlap—resulting in a purer trapped charge.

(b) the lower thermal loading associated with engines in which pistons, cylinder heads and liners are exposed to the most severe pressures and temperatures associated with combustion only every other revolution.

(c) Easier lubrication conditions for pistons, rings and liners due to the absence of ports, and the idle stroke renewing liner lubrication and giving inertia lift off to rings and small and large end bearings.

These factors make it possible for the four-stroke engine to achieve output levels of the order of 75% of equivalent two stroke engines. In recent years attention has focused particularly on three-cylinder high speed passenger car two-stroke engines as a possible replacement for conventional four-cylinder, four stroke engines with considerable potential savings in space and weight.

In engines of this type admission of air is effected by ‘air piston’ controlled inlet ports, and rejection of products of combustion by ‘exhaust piston’ controlled exhaust ports. The motion of the two sets of pistons is controlled by either two crankshafts connected through gearing, or by a single crankshaft with the ‘top’ bank of pistons transmitting their motion to the single crankshaft through a crosshead-siderod mechanism. By suitable offsetting of the cranks controlling the air and exhaust pistons asymmetrical timing can be achieved.

It is evident that this system displays the same favourable characteristics as the exhaust valve in head system, but at the expense of even greater mechanical complications. Its outstanding advantage is the high specific output per cylinder associated with two pistons. However, the system is now retained only in large low speed marine, and smaller medium speed stationary and marine engines. In high speed form it is still employed for naval purposes such as in some fast patrol vessels and mine searchers, although its use in road vehicles and locomotives is discontinued.

In engines of this type admission of air to the cylinder is usually effected by piston controlled ports while the products of combustion are exhausted through a camshaft operated exhaust valve. Such systems are preferable from the standpoint of scavenging in that the ‘uniflow’ motion of the air from the inlet ports upwards through the cylinder tends to lead to physical displacement of, rather than mixing with, the products of combustion thus giving improved charge purity at the end of the scavenging process. At the same time it is now possible to adopt asymmetrical timing of the exhaust and inlet processes relative to bottom dead centre (BDC) so that, with exhaust closure preceding inlet closure the danger of escape of fresh charge into the exhaust manifold present in the loop scavenge system is completely eliminated. This system has been adopted in a number of stationary and marine two-stroke engines.